U.S. patent number 10,156,163 [Application Number 15/280,577] was granted by the patent office on 2018-12-18 for system and method for variable actuation of a valve of an internal-combustion engine, with a device for dampening pressure oscillations.
This patent grant is currently assigned to C.R.F. SOCIETA CONSORTILE PER AZIONI. The grantee listed for this patent is C.R.F. Societa Consortile per Azioni. Invention is credited to Onofrio De Michele, Marcello Gargano, Carlo Mazzarella, Raffaele Ricco, Sergio Stucchi.
United States Patent |
10,156,163 |
Stucchi , et al. |
December 18, 2018 |
System and method for variable actuation of a valve of an
internal-combustion engine, with a device for dampening pressure
oscillations
Abstract
A system for variable actuation of an engine valve of an
internal-combustion engine includes a master piston and a slave
piston, driven by the master piston. A control valve controls a
communication of a volume of a pressurized fluid with a lower
pressure environment, which is connected to a fluid accumulator,
and an electronic control unit controls the electrically operated
control valve. A device for dampening pressure oscillations is
connected to the volume of pressurized fluid and includes an
additional volume adapted for receiving fluid from the volume of
pressurized fluid only when, following upon oscillations of the
pressure in the volume of pressurized fluid, the pressure exceeds a
maximum threshold value, which is higher than a mean pressure value
that is set up in the volume of pressurized fluid when the master
piston drives the slave piston in normal operating conditions.
Inventors: |
Stucchi; Sergio (Valenzano,
IT), Ricco; Raffaele (Casamassima, IT),
Gargano; Marcello (Torre a Mare, IT), De Michele;
Onofrio (Castellana Grotte, IT), Mazzarella;
Carlo (Noicattaro, IT) |
Applicant: |
Name |
City |
State |
Country |
Type |
C.R.F. Societa Consortile per Azioni |
Orbassano (Turin) |
N/A |
IT |
|
|
Assignee: |
C.R.F. SOCIETA CONSORTILE PER
AZIONI (Orbassano (Turin), IT)
|
Family
ID: |
54697432 |
Appl.
No.: |
15/280,577 |
Filed: |
September 29, 2016 |
Prior Publication Data
|
|
|
|
Document
Identifier |
Publication Date |
|
US 20170101903 A1 |
Apr 13, 2017 |
|
Foreign Application Priority Data
|
|
|
|
|
Oct 13, 2015 [EP] |
|
|
15189506 |
|
Current U.S.
Class: |
1/1 |
Current CPC
Class: |
F02D
13/0207 (20130101); F01L 1/25 (20130101); F01L
9/025 (20130101); F01L 2810/04 (20130101); F01L
2810/03 (20130101); F01L 2001/0537 (20130101) |
Current International
Class: |
F01L
1/25 (20060101); F02D 13/02 (20060101); F01L
9/02 (20060101); F01L 1/053 (20060101) |
Field of
Search: |
;123/90.12 |
References Cited
[Referenced By]
U.S. Patent Documents
Foreign Patent Documents
|
|
|
|
|
|
|
2926327 |
|
Jan 1981 |
|
DE |
|
3625664 |
|
Feb 1988 |
|
DE |
|
102005000621 |
|
Jul 2006 |
|
DE |
|
0498682 |
|
Jan 1992 |
|
EP |
|
0803642 |
|
Nov 2000 |
|
EP |
|
1091097 |
|
Apr 2001 |
|
EP |
|
1243761 |
|
Sep 2002 |
|
EP |
|
1344900 |
|
Sep 2003 |
|
EP |
|
1508676 |
|
Feb 2005 |
|
EP |
|
1555398 |
|
Jul 2005 |
|
EP |
|
1674673 |
|
Jun 2006 |
|
EP |
|
1726790 |
|
Nov 2006 |
|
EP |
|
2597276 |
|
May 2013 |
|
EP |
|
2806195 |
|
Nov 2014 |
|
EP |
|
2070716 |
|
Sep 1981 |
|
GB |
|
Other References
Corresponding European Search Report for EP application No.
15189506, completed on Mar. 17, 2016, and dated Mar. 24, 2016.
cited by applicant.
|
Primary Examiner: Newton; Jason
Attorney, Agent or Firm: Cardona; Victor Heslin Rothenberg
Farley & Mesiti P.C.
Claims
What is claimed is:
1. A system for variable actuation of an engine valve of an
internal-combustion engine, comprising: a master piston driven
directly or indirectly by a cam of a camshaft of the
internal-combustion engine; a slave piston, which drives said
engine valve and is hydraulically driven by said master piston, by
means of a volume of pressurized fluid interposed between the
master piston and the slave piston; an electrically operated
control valve, which controls a communication of said volume of
pressurized fluid with a lower pressure environment, said lower
pressure environment being connected to a fluid accumulator; and an
electronic control unit for controlling said electrically operated
control valve on the basis of one or more parameters indicating the
operating conditions of the engine and/or of the system for
variable actuation of the engine valve, a device for dampening
pressure oscillations in the volume of pressurized fluid connected
to said volume of pressurized fluid, and said oscillation dampening
device comprising an additional volume adapted for receiving fluid
from said volume of pressurized fluid only when said pressure
exceeds a maximum threshold value.
2. The system according to claim 1, further comprising an auxiliary
chamber comprising said additional volume, said auxiliary chamber
in communication with said volume of pressurized fluid and defined
by a movement of a movable member against the action of a return
spring, said spring having a load such that said movable member
displaces against the action of the spring, thus creating said
additional volume, only when the pressure in the volume of
pressurized fluid exceeds the aforesaid maximum threshold
value.
3. The system according to claim 2, wherein the communication of
the aforesaid auxiliary chamber with the volume of pressurized
fluid is a permanently opened communication.
4. The system according to claim 3, wherein said permanently opened
communication includes a restricted passage.
5. The system according to claim 2, wherein said auxiliary chamber
and said movable member are provided within the body of the
oscillation dampening member, said body comprising a separate
element.
6. The system according to claim 2, wherein said auxiliary chamber
and said movable member are provided within the body of said slave
piston.
7. The system according to claim 2, wherein said auxiliary chamber
and said movable member are provided within the body of said master
piston.
8. The system according to claim 2, wherein said auxiliary chamber
and said movable member are provided within the body of said
electrically operated control valve.
9. The system according to claim 1, further comprising an auxiliary
chamber in communication with said volume of pressurized fluid and
defined by a movement of a movable member, the position of the
movable member being controlled by an electrically driven actuator,
said electronic control unit being programmed for controlling said
actuator so as to cause a displacement thereof which creates the
aforesaid additional volume when the pressure in the volume of
pressurized fluid exceeds said maximum threshold value.
10. The system according to claim 9, wherein the threshold pressure
value that triggers the actuator is varied as a function of the
operating conditions.
11. The system according to claim 9, wherein said electronic
control unit controls the actuator of said oscillation dampening
device in a closed-loop mode, on the basis of a signal from at
least one pressure sensor adapted to detect the pressure in the
volume of pressurized fluid.
12. The system according to claim 9, wherein said electronic
control unit is programmed for controlling the actuator of the
movable member of the oscillation dampening device in an open-loop
mode, on the basis of stored maps, as a function of the operating
conditions of the engine and/or of the system for variable
actuation of the engine valve.
13. The system according to claim 9, wherein the communication of
the aforesaid auxiliary chamber with the volume of pressurized
fluid is a permanently opened communication.
14. The system according to claim 9, wherein said auxiliary chamber
and said movable member are provided within the body of the
oscillation dampening member, said body comprising a separate
element.
15. The system according to claim 9, wherein said auxiliary chamber
and said movable member are provided within the body of said slave
piston.
16. The system according to claim 9, wherein said auxiliary chamber
and said movable member are provided within the body of said master
piston.
17. The system according to claim 9, wherein said auxiliary chamber
and said movable member are provided within the body of said
electrically operated control valve.
Description
CROSS REFERENCE TO RELATED APPLICATIONS
This application claims priority from European patent application
No. 15189506.7 filed on Oct. 13, 2015, the entire disclosure of
which is incorporated herein by reference.
FIELD OF THE INVENTION
The present invention relates to systems for variable actuation of
engine valves for internal-combustion engines, of the type
comprising:
a master piston driven directly or indirectly by a cam of a
camshaft of the internal-combustion engine;
a slave piston, which drives said engine valve and is hydraulically
driven by said master piston, by means of a volume of pressurized
fluid interposed between the master piston and the slave
piston;
an electrically operated control valve, which controls a
communication of said volume of pressurized fluid with a lower
pressure environment, said lower pressure environment being
connected to a fluid accumulator; and
an electronic control unit for controlling said electrically
operated control valve on the basis of one or more parameters
indicating the operating conditions of the engine and/or of the
system for variable actuation of the engine valves.
PRIOR ART
Since long, the present applicant has been developing
internal-combustion engines provided with a system of the above
indicated type, for variable actuation of the intake valves,
marketed under the trademark "Multiair", this system having the
features referred to above. The present applicant is the assignee
of many patents and patent applications relating to engines
provided with a system of this type and to components of this
system.
FIG. 1 of the annexed drawings shows a cross-sectional view of a
cylinder head of an internal-combustion engine according to the
technique described in EP 0 803 642 B1. The cylinder head
illustrated in FIG. 1 and designated by the reference number 1 is
applied to an engine with four cylinders in line; however, the
variable-actuation system illustrated therein is of general
application. The cylinder head 1 comprises, for each cylinder, a
cavity 2, which is formed in the base surface 3 of the cylinder
head 1 and defines the combustion chamber. Giving out into the
cavity 2 are two intake ducts 4, 5 (the duct 5 is represented with
a dashed line) and two exhaust ducts 6 (only one of which is
visible in the figure). Communication of the two intake ducts 4, 5
with the combustion chamber 2 is controlled by two intake valves 7
(only one of which is visible in the figure), of the traditional
poppet type, each comprising a stem 8 slidably mounted in the body
of the cylinder head 1.
Each valve 7 is recalled into the closing position by springs 9
interposed between an internal surface of the cylinder head 1 and
an end valve retainer 10. Communication of the two exhaust ducts 6
with the combustion chamber is controlled by two valves 70 (only
one of which is visible in the figure), which are also of a
conventional type and associated to which are springs 9 for return
towards the closed position.
Opening of each intake valve 7 is controlled, in the way that will
be described in what follows, by a camshaft 11, which is rotatably
mounted about an axis 12 within supports of the cylinder head 1 and
comprises a plurality of cams 14 for actuation of the intake valves
7 of the internal-combustion engine.
Each cam 14 that controls an intake valve 7 co-operates with the
plate 15 of a tappet 16 slidably mounted along an axis 17, which,
in the case of the example illustrated in the prior document cited,
is set substantially at 90.degree. with respect to the axis of the
valve 7. The plate 15 is recalled against the cam 14 by a spring
associated thereto. The tappet 16 constitutes a pumping plunger, or
master piston, slidably mounted within a bushing 18 carried by a
body 19 of a pre-assembled unit 20, which incorporates all the
electrical and hydraulic devices associated to actuation of the
intake valves, according to what is described in detail in what
follows. There may be provided a separate unit 20 for each cylinder
of the engine.
The master piston 16 is able to transmit a force to the stem 8 of
the valve 7 so as to cause opening of the latter against the action
of the elastic means 9, by means of pressurized fluid (preferably
oil coming from the engine-lubrication circuit) present in a volume
of pressurized fluid C facing which is the master piston 16, and by
means of a slave piston 21 slidably mounted in a cylindrical body
constituted by a bushing 22, which is also carried by the body 19
of the pre-assembled unit 20.
Once again with reference to FIG. 1, the volume of pressurized
fluid C associated to each intake valve 7 can be set in
communication with a lower pressure environment, constituted by an
exhaust channel 23, via a solenoid valve 24. The channel 23 is
designed to receive from the engine-lubrication circuit oil
supplied by the pump of the lubrication circuit, via a duct
arranged in which are one or more bleeding siphons and a non-return
valve (see in this connection, for example, EP-A-1 243 761 and
EP-A-1 555 398 in the name of the present applicant).
The solenoid valve 24, which may be of any known type suitable for
the purpose illustrated herein, is controlled by electronic control
means 25, as a function of signals S indicating operating
parameters of the engine, such as the position of the accelerator
and the engine r.p.m. or the temperature or viscosity of the oil in
the system for variable actuation of the valves.
When the solenoid of the solenoid valve 24 is energized, the
solenoid valve is closed so as to maintain the volume of fluid C
under pressure and enable actuation of each intake valve 7 by the
respective cam 14, via the master piston 16, the slave piston 21,
and the volume of oil comprised between them.
When the solenoid of the solenoid valve 24 is de-energized, the
solenoid valve opens so that the volume C enters into communication
with the channel 23, and the pressurized fluid present in the
volume C flows into this channel. Consequently, a decoupling is
obtained of the cam 14 and of the master piston 16 from the intake
valve 7, which thus returns rapidly into its closing position under
the action of the return springs 9.
By controlling the communication between the volume C and the
exhaust channel 23, it is consequently possible to vary the timing
of opening and/or closing and the opening lift of each intake valve
7.
The exhaust channels 23 of the various solenoid valves 24 all give
out into one and the same longitudinal channel 26 communicating
with pressure accumulators 27, only one of which is visible in FIG.
1. Each accumulator is substantially constituted by a cylindrical
body in which a plunger is slidably mounted, defining an
accumulator chamber, which communicates with the low-pressure
environment defined by the exhaust channels 23, 26. A helical
spring within the accumulator recalls the plunger of the
accumulator into a position in which the volume for receiving the
fluid within the accumulator is minimum. When the solenoid valve 24
is opened, part of the pressurized fluid coming from the volume C
flows into the accumulator 270.
The master piston 16 with the associated bushing 18, the slave
piston 21 with the associated bushing 22, the solenoid valve 24,
and the channels 23, 26 are carried by, or formed in, the aforesaid
body 19 of the pre-assembled unit 20, to the advantage of rapidity
and ease of assembly of the engine.
In the example illustrated, the exhaust valves 70 associated to
each cylinder are controlled in a conventional way, by a respective
camshaft 28, via respective tappets 29, even though in principle
there is not excluded application of the variable-actuation system
also to the exhaust valves. This applies also to the present
invention.
Once again with reference to FIG. 1, the variable-volume chamber
defined inside the bushing 22 and facing the slave piston 21 (which
in FIG. 1 is illustrated in its condition of minimum volume, given
that the slave piston 21 is at its top dead centre) communicates
with the pressurized-fluid chamber C via an opening 30 made in an
end wall of the bushing 22. This opening 30 is engaged by an end
nose 31 of the plunger 21 in such a way as to provide hydraulic
braking of the movement of the valve 7 in the closing phase, when
the valve is close to the closing position, in so far as the oil
present in the variable-volume chamber is forced to flow into the
volume of pressurized fluid C passing through the clearance
existing between the end nose 31 and the wall of the opening 30
engaged thereby. In addition to the communication constituted by
the opening 30, the volume of pressurized fluid C and the
variable-volume chamber of the slave piston 21 communicate with one
another via internal passages made in the body of the slave piston
21 and controlled by a non-return valve 32, which enables passage
of fluid only from the pressurized volume C to the variable-volume
chamber of the slave piston 21. Various alternative embodiments of
the hydraulic-braking device of the slave piston 21 have been
proposed in the past by the present applicant (see, for example,
EP-A-1 091 097 and EP-A-1 344 900). The purpose of the
hydraulic-braking device is to prevent a sharp impact (and
consequent noise) of the valve 7 against its seat when the valve 7
returns rapidly into the closing position following upon opening of
the solenoid valve 24.
During normal operation of the known engine illustrated in FIG. 1,
when the solenoid valve 24 excludes communication of the volume of
pressurized fluid C with the exhaust channel 23, the oil present in
the volume C transmits the movement of the master piston 16,
imparted by the cam 14, to the slave piston 21, which drives
opening of the valve 7. In the reverse movement of closing of the
valve, as has already been said, during the final step the nose 31
enters the opening 30 causing hydraulic braking of the valve so as
to prevent impact of the body of the valve against its seat, for
example following upon an opening of the solenoid valve 24 that
causes immediate return of the valve 7 into the closing
position.
In the system described, when the solenoid valve 24 is activated,
the engine valve follows the movement of the cam (full lift). An
early closing of the valve can be obtained by opening the solenoid
valve 24 so as to empty out the volume of pressurized fluid C and
obtain closing of the valve 7 under the action of the respective
return springs 9. Likewise, a late opening of the valve can be
obtained by delaying closing of the solenoid valve, whereas the
combination of a late opening and an early closing of the valve can
be obtained by closing and opening the solenoid valve during the
thrust of the corresponding cam. According to an alternative
strategy, in line with the teachings of EP 1 726 790 A1 in the name
of the present applicant, each intake valve can be controlled in a
"multi-lift" mode, i.e., according to two or more repeated
"sub-cycles" of opening and closing. In each subcycle, the intake
valve opens and then closes completely. The electronic control unit
is consequently able to obtain a variation of the timing of opening
and/or closing and/or of the lift of the intake valve, as a
function of one or more operating parameters of the engine. This
enables the maximum engine efficiency to be obtained, and the
lowest fuel consumption, in every operating condition.
FIG. 2 of the annexed drawings corresponds to FIG. 6 of EP 1 674
673 in the name of the present applicant and shows a diagram of the
system for actuation of the two intake valves associated to each
cylinder, in a conventional Multiair system. This figure shows two
intake valves 7 associated to one and the same cylinder of an
internal-combustion engine, which are controlled by a single master
piston 16, which is in turn controlled by a single cam of the
engine camshaft (not illustrated) acting against a plate 15. FIG. 2
does not illustrate the return springs 9 (see FIG. 1) that are
associated to the valves 7 and tend to bring them back into the
respective closed positions. As may be seen, in the conventional
system of FIG. 2, a single master piston 16 controls the two intake
valves 7 via a single volume of pressurized fluid C, the
communication with discharge being controlled by a single solenoid
valve 24. The volume of pressurized fluid C is in hydraulic
communication with both of the variable-volume chambers C1, C2
facing two slave pistons 21 for control of the intake valves 7 of
one and the same cylinder.
The system of FIG. 2 is able to operate in an efficient and
reliable way above all in the case where the volumes of the
hydraulic chambers are relatively small. This possibility is
afforded by adopting hydraulic tappets 400 on the outside of the
bushings 22, according to what has already been illustrated in
detail, for example, in EP 1 674 673 B1 in the name of the present
applicant. In this way, the bushings 22 may have an internal
diameter that can be chosen as small as desired.
FIG. 3 of the annexed drawings is a schematic representation of the
system illustrated in FIG. 2, in which it is evident that both of
the intake valves 7 associated to each cylinder of the engine have
the hydraulic chambers of the two slave pistons 21 permanently in
communication with the pressurized volume C, which in turn may be
isolated or connected to the exhaust channel 23, via the single
solenoid valve 24.
The solution illustrated in FIGS. 2 and 3 enables obvious
advantages as regards simplicity and economy of construction, and
from the standpoint of reduction of the overall dimensions, as
compared to the solution illustrated, for example, in EP 0 803 642
B1, which envisages two solenoid valves for controlling separately
the two intake valves of each cylinder.
On the other hand, the solution with a single solenoid valve per
cylinder rules out the possibility of differentiating control of
the intake valves of each cylinder. This differentiation is,
instead, desired, in particular in the case of diesel engines where
each cylinder is provided with two intake valves associated to
respective intake ducts having different shapes from one another in
order to generate different movements of the flow of air introduced
into the cylinder (see, for example, FIG. 5 of EP 1 508 676 B1).
Typically, in these engines the two intake ducts of each cylinder
are shaped for obtaining optimized TUMBLE-type and SWIRL-type flows
of air, respectively, these flow types being fundamental for
optimal distribution of the charge of air within the cylinder,
which greatly affects the possibility of reducing the pollutant
emissions at the exhaust.
In order to solve the above problem, the present applicant has also
proposed the use of a different system layout, which makes use of a
three-position and three-way solenoid valve, as described for
example in EP 2 597 276 A1 in the name of the present
applicant.
Once again with reference to the known systems to which the present
invention can be applied, the present applicant has proposed in the
past also alternative solutions for the electrically operated
control valve 24, which may be, instead of a solenoid valve, an
electrically operated valve of any other type, for example a valve
with a piezoelectric actuator or a magnetostrictive actuator (EP 2
806 195 A1).
For the purposes of application of the present invention, all the
variants described above may likewise be adopted.
FIG. 3A of the annexed drawings shows a perspective view of the
main components of a known embodiment of the Multiair system of the
present applicant (the components associated to one cylinder of the
engine are shown), corresponding to the general scheme of FIGS. 2
and 3 of the annexed drawings. In FIG. 3A, the parts corresponding
to those of FIGS. 1-3 are designated by the same reference
numbers.
In the case of the embodiment of FIG. 3A, the master piston 16 is
driven by the respective cam 14 via a rocker arm 140 having an
intermediate portion carrying a freely rotatable roller 141
engaging with the cam 14. The rocker arm 140 has one end rotatably
supported by a supporting element 142 mounted in the pre-assembled
unit 20. The opposite end of the rocker arm 140 engages with the
plate 15 of the master piston 16. FIG. 3A does not show the spring
that recalls the plate 15 against the cam 14. FIG. 3A shows the
communications of the high-pressure volume C with the solenoid
valve 24 and the solenoid valve 24 with the chambers associated to
the two slave pistons 21.
Technical Problem
Studies and tests conducted by the present applicant have shown
that in given operating conditions the systems for variable
actuation of the valves of the type indicated above are subject to
pressure oscillations inside the high-pressure volume. These
pressure oscillations are due to the movement of the master piston,
which pressurizes the oil present in the high-pressure volume with
a dynamics that depends upon various operating factors, such as the
type of movement of the master piston (linked to the profile of the
cam), the specific operating condition of the system, the size of
the high-pressure volume. Pressure oscillations occur in
particular, for example, in the "Late Intake Valve Opening" (LIVO)
mode, i.e., when opening of the intake valve is delayed with
respect to the conventional cycle determined by the profile of the
cam, through a delayed closing of the electrically operated control
valve.
Pressure oscillations in the high-pressure volume introduce various
disadvantages, amongst which in particular noise and vibrations and
a shorter service life of the components of the system.
Object of the Invention
The object of the present invention is to provide a system for
variable actuation of the valves of an internal-combustion engine
that will be able to overcome the drawback indicated above.
A further object of the invention is to achieve the above purpose
by adopting means that are simple, low-cost, and safe and reliable
in operation.
SUMMARY OF THE INVENTION
With a view to achieving the aforesaid objects, the subject of the
present invention is a system for variable actuation of an engine
valve of an internal-combustion engine, comprising:
a master piston driven directly or indirectly by a cam of a
camshaft of the internal-combustion engine;
a slave piston, which drives said engine valve and is hydraulically
driven by said master piston, by means of a volume of pressurized
fluid interposed between the master piston and the slave
piston;
an electrically operated control valve, which controls a
communication of said volume of pressurized fluid with a lower
pressure environment, which is connected to a fluid accumulator;
and
an electronic control unit for controlling said electrically
operated control valve on the basis of one or more parameters
indicating the operating conditions of the engine and/or of the
system for variable actuation of the engine valves,
said oscillation dampening device comprising an additional volume
adapted for receiving fluid from said volume of pressurized fluid
only when said pressure exceeds a maximum threshold value.
Tests conducted by the present applicant have shown that, thanks to
the aforesaid characteristics, the problem of pressure oscillations
in the volume of fluid at high pressure is solved in a simple and
efficient way, substantially reducing the vibrations and noise of
the system and consequently enabling a longer service life of its
components.
The invention may be applied to any type of system for variable
actuation of the engine valves of the type comprising a master
piston, a slave piston, and a volume of pressurized fluid
interposed between them that can be connected with a low-pressure
environment for decoupling the engine valve from the actuation cam.
The invention may be applied irrespective of the architecture of
the system (with one electrically operated control valve or with
two electrically operated control valves for control of the two
intake valves of one and the same cylinder, and with electrically
operated valves of a normally open type or a normally closed type).
The electrically operated valve may be of the two-way, two-position
type, or of the three-way, three-position type, or of any other
type and may envisage actuation by means of a solenoid or else any
other type of actuator (for example a piezoelectric or
magnetostrictive actuator). The invention may also apply to systems
for variable actuation of the engine exhaust valves.
In a first embodiment of the invention, the aforesaid additional
volume is constituted by an auxiliary chamber that is in
communication with the above volume of pressurized fluid and is
defined by the movement of a movable member against the action of a
return spring, the spring having a load such that the movable
member displaces against the action of the spring, thus creating
the additional volume only when the pressure in the volume of
pressurized fluid exceeds the aforesaid maximum threshold
value.
In the above mentioned first embodiment, the oscillation dampening
device operates automatically whenever in the high-pressure volume
a pressure peak above the maximum threshold value is generated.
In a second embodiment of the invention, the additional volume is
constituted by an auxiliary chamber that is in communication with
the volume of pressurized fluid and is defined by movement of a
movable member the position of which is controlled by an
electrically driven actuator, the electronic control unit being
programmed for controlling the actuator so as to cause displacement
thereof and thus create the aforesaid additional volume when the
pressure in the volume of pressurized fluid exceeds the above
maximum threshold value.
In this second embodiment, the electronic control unit controls in
closed-loop mode the aforesaid actuator of the oscillation
dampening device on the basis of the signal at output from at least
one pressure sensor that is designed to detect the pressure in the
volume of pressurized fluid, or else is programmed for operating in
open-loop mode, on the basis of stored maps, as a function of the
operating conditions of the engine and/or of the system for
variable actuation of the engine valves.
The advantage of this second embodiment lies in the fact that the
triggering pressure threshold is not fixed as in the solution with
automatic operation, but can be varied as a function of the
operating conditions. Moreover, the actuator associated to the
damper device may be either of the on/off type or of a proportional
type.
In both of the aforesaid embodiments, the communication of the
auxiliary chamber with the volume of pressurized may be a
permanently opened communication, which preferably includes a
restricted passage in order to isolate the high-pressure volume in
regard to possible pressure oscillations within the aforesaid
auxiliary chamber of the device for dampening oscillations.
In a first solution, the aforesaid auxiliary chamber and the
aforesaid movable member of the oscillation dampening device are
provided within the body of an autonomous member, associated to the
high-pressure volume. In variants of said solution, the auxiliary
chamber and the movable member are provided within the body of the
slave piston, or within the body of the master piston, or within
the body of the electrically operated control valve.
BRIEF DESCRIPTION OF THE DRAWINGS
Further characteristics and advantages of the invention will emerge
from the ensuing description with reference to the annexed
drawings, which are provided purely by way of non-limiting example
and in which:
FIG. 1 is a cross-sectional view of a cylinder head of an
internal-combustion engine provided with a system for variable
actuation of the intake valves according to the known art;
FIG. 2 is a diagram of a system for variable actuation of the
valves of an internal-combustion engine according to the known
art;
FIG. 3 is a further diagram of the system of FIG. 2;
FIG. 3A is a perspective view of an embodiment of the known system
represented schematically in FIGS. 2 and 3;
FIG. 4 is a diagram similar to that of FIG. 3 that shows the basic
principle of the system according to the invention;
FIG. 4A shows the same embodiment of FIG. 3A, modified according to
the present invention;
FIGS. 5A and 5B are cross-sectional views that show the member for
dampening oscillations forming part of the solution of FIG. 4A, in
two different operating conditions;
FIGS. 6 and 7 are plots that show the substantial reduction and/or
elimination of the pressure oscillations in the high-pressure
volume, which can be obtained with the system according to the
invention;
FIG. 8 shows a variant of FIG. 4;
FIG. 9 is a cross-sectional view of an embodiment of a oscillation
dampening device that can be used in the system of FIG. 8;
FIG. 10 is a cross-sectional view of the slave-piston assembly of
the system of FIG. 8, which incorporates the oscillation dampening
device of FIG. 9;
FIG. 11 shows a further variant of the system of FIG. 4;
FIG. 11A illustrates an embodiment of the member for dampening
oscillations incorporated in the master piston of the system of
FIG. 11;
FIG. 12 shows a further variant of the system of FIG. 4; and
FIG. 13 is a diagram of a further embodiment of the system
according to the invention that uses a oscillation dampening device
with controlled triggering.
DETAILED DESCRIPTION OF SOME PREFERRED EMBODIMENTS
FIGS. 1-3 and 3A, which relate to the prior art, have already been
described above. FIG. 4 of the annexed drawings is a schematic view
similar to that of FIG. 3 and regards the system for variable
actuation of the valves of an internal-combustion engine according
to the present invention. In FIG. 4, the parts corresponding to
those of FIG. 3 are designated by the same reference numbers. The
main difference of the system according to the invention as
illustrated in FIG. 4 as compared to the known system of FIG. 3
lies in the fact that the high-pressure volume C is connected to a
device for damping pressure oscillations D. A recirculation line
800 connects the rear side of the device D with the low-pressure
line 23, or with the accumulator 270, according to what will be
illustrated in detail in what follows.
FIGS. 4A, 5A, and 5B show a first example of embodiment of the
system according to the invention, in which the dampening
oscillation device D is constituted as autonomous member associated
to the high-pressure volume C. One damper device D is provided for
each cylinder of the engine. FIG. 4A shows the same perspective
view as that of FIG. 3A, modified according to the teachings of the
present invention. As will be seen, in the example illustrated in
FIG. 4A, the device D is directly associated to a channel for
communication between the chamber of the master piston and the
solenoid valve 24, this channel forming part of the high-pressure
volume C. The damper device D of this embodiment is illustrated in
cross-sectional view and at an enlarged scale in FIGS. 5A, and 5B,
in two different operating conditions. The device D of this
embodiment is received in a corresponding seat formed in the
pre-assembled unit 20 already described above, which carries all
the elements of the system for variable actuation of the engine
valves. As already mentioned there may be provided a separate unit
20 for each cylinder.
With reference to FIGS. 5A and 5B, which are provided purely by way
of non-limiting example, the oscillation dampening deviceD of this
embodiment comprises a cylindrical body D1 having an internal
cylindrical cavity D2 slidably mounted within which is a movable
member D3. A helical spring D4 is interposed axially between the
movable member D3 and a bushing D5 received and blocked within the
cylindrical cavity of the body D1 with interposition of seal rings
D6. The helical spring D4 tends to maintain the movable member D3
in an end-of-travel position, in the direction of a chamber D7,
which is defined within the cylindrical body D1 and communicates
with a hole D8 of an end connector D9, designed to be set in
hydraulic connection with the high-pressure volume C, as may be
seen in FIG. 4A. The chamber D7 communicates with the hole D8 of
the connector D9 via a restricted passage D10 of a predetermined
diameter, formed in the bottom wall of a cup-shaped element D11
that is secured, by being driven into the cylindrical body D1 or
with a threaded connection, against a bottom wall of the internal
cavity D2 of the device, in which the aforesaid hole D8 gives out.
In the example illustrated, the movable member D3 has a cup-shaped
body with a bottom wall facing the chamber D7 and an internal
cavity D31 that faces the spring D4 and is in communication with
the low-pressure environment of the circuit through the internal
cavity D51 of the bushing D5, the end portion of the internal
cavity D2 of the body D1, and the recirculation line 800 (see FIG.
4). Once again in the case of the specific example illustrated,
axially interposed between the spring D4 and the bushing D5 is a
ring D12.
As already mentioned, the oscillation dampening deviceD is
prearranged in such a way that the chamber D7 is permanently in
communication, via the restricted passage D10 and the hole D8 of
the connector D9, with the high-pressure volume C associated to a
cylinder of the engine.
FIG. 5A shows the device D in the inactive resting condition, in
which the spring D4 maintains the plunger D3 in an end-of-travel
position, against an annular contrast portion formed in the
internal cavity D2 of the body D1. In this condition, the volume
internal to the device D that is in communication with the
high-pressure volume of the system for actuation of the valves is
substantially that of the chamber D7, defined within the cup-shaped
element D10 and limited at the top by the movable member D3, held
in its resting position (the lowest position, as viewed in the
drawings). The volume internal to the device D further comprises
the restricted hole D11 and the duct D8. This internal volume is
always filled with fluid during normal operation of the system for
variable actuation of the engine valves, being permanently in
communication with the high-pressure volume C of the system.
During operation of the system for variable actuation of the engine
valves, in the case where the pressure of the fluid in the
high-pressure volume C presents oscillations with peaks higher than
a predetermined threshold value, markedly higher than the mean
value of the pressure that is set up in the volume C during normal
driving of the slave pistons 21 by the master piston 16, these
pressure peaks manage to overcome the action of the spring D4,
causing displacement of the movable member D3 against the spring D4
and consequent formation within the cavity D2 of the device D of an
additional volume D7' formed between the annular contrast portion
of the cavity D2 that defines the resting position of the movable
member D3 and the surface of the movable member facing it. In other
words, this additional volume basically corresponds to the portion
of the internal cavity D2 that is left free by the movable member
D3 when this moves away from the resting position illustrated in
FIG. 5A so as to move into the operating position of FIG. 5B.
The characteristics of the spring D4 and the loading of the spring
in its resting position (which may also be varied using rings D12
of a different height) are predetermined in such a way that the
pressure of fluid that is able of cause displacement of the movable
member D3 is a threshold value notably higher than the mean
pressure value that is set up in the high-pressure volume C when
the master piston controls each slave piston 21 in normal operating
conditions. Consequently, the damper device D enters into action
only when the pressure in the volume C has anomalous oscillations
and consequent pressure peaks above the threshold value.
Moreover, sizing of the device D is chosen in such a way that the
additional volume D7' that is created in the case of pressure peaks
is the one necessary and sufficient for dampening the pressure
oscillations and does not appreciably alter the desired stroke of
the slave pistons 21 caused by the movement of the master
piston.
Purely by way of example, the additional volume D7' that is set up
in the case of pressure peaks corresponds to approximately 1% of
the total high-pressure volume C associated to each cylinder of the
engine.
In summary, the damper device according to the invention is able to
increase the overall volume of the high-pressure environment
whenever there arise pressure peaks, thus attenuating the pressure
oscillations accordingly. Dampening of the oscillations produces
the beneficial effect of reducing drastically or even eliminating
altogether vibrations and noise of the system, with consequent
advantage also as regards the service life of the components of the
system. For operation of the system, it is necessary for the damper
device D to "see" always the high-pressure volume C in which the
pressure oscillation is to be attenuated.
The embodiment of FIGS. 5A and 5B is characterized in that it
entails an automatic triggering of the damper whenever the pressure
exceeds the threshold value defined above, for which the spring D4
is provided.
It is possible to pre-determine the increase in volume D7' that is
necessary, knowing the amplitude of the pressure oscillations that
are to be attenuated and sizing accordingly the diameter of the
movable member and adopting a spring having the necessary
stiffness.
The restricted passage D10 has the function of filtering the
pressure oscillations that are generated within the damper device
D, preventing propagation thereof into the high-pressure volume
C.
A dynamic seal between the body D2 of the device and the movable
member D3 may be obtained by means of an adequate control of the
coupling clearance, thus allowing a minimum leakage of fluid
towards the low-pressure environment through the recirculation line
800, or else by pre-arrangement of dynamic seals, made, for
example, of plastic material, which are designed to prevent
leakage. In any case, when the plunger is in its end-of-travel
position in the direction of the spring D4, it comes into contact
with an end surface of the bushing D5, closing communication with
the hole D51.
FIG. 6 shows by way of example the attenuation of the pressure
oscillations that can be obtained with an oscillation dampening
device D of the type illustrated in FIGS. 5A and 5B.
In FIG. 6, the plot represented with a dashed line indicates the
variation of pressure in the high-pressure volume C as a function
of the crank angle in a system according to the known art, i.e.,
without the damper device.
The plot of FIG. 6 shows an example of embodiment in "LIVO" mode in
which the intake valve opens with a delay with respect to what
would be obtained by the cam profile. In the case illustrated, the
increase in pressure that causes opening of the intake valve is in
fact at a crank angle of approximately 450.degree., i.e.,
substantially at half of the descent of the engine piston from top
dead centre (360.degree.) to bottom dead centre (540.degree.). As
may be seen, in the operating step of actuation of the engine
valve, in which the solenoid valve 24 is closed for pressurizing
the volume C and enabling the master cylinder 16 to drive via the
volume of pressurized oil displacement of each slave piston 21, the
pressure presents rather significant oscillations around its mean
value, with pressure peaks well above the aforesaid mean value.
The plot represented with a solid line in FIG. 6 shows the
corresponding variation of the pressure in the volume C in the case
of a system provided with the oscillation dampening device of the
type of FIGS. 5A and 5B. As may be seen, all other conditions being
the same, the pressure oscillations in the volume C are markedly
attenuated.
FIG. 7 shows the frequency response regarding the variation of the
pressure in the high-pressure volume, respectively in the case of
the known system, without the oscillation dampening device (dashed
line), and in the case of the system provided with a oscillation
dampening device according to the invention. It may be noted that,
in the example considered herein, at the lower frequencies there is
a considerable reduction of the amplitude of the pressure
oscillations.
FIG. 8 is a schematic illustration of a variant of the system
according to the invention, where the oscillation dampening device
D is associated to one (or possibly to each) of the two slave
pistons 21.
FIGS. 9 and 10 refer to an example of embodiment of this variant.
FIG. 10 is a cross-sectional view at an enlarged scale of a slave
piston 21, according to a known embodiment of the Multiair system,
here modified for receiving the oscillation dampening device D.
FIG. 9 shows the oscillation dampening device D just by itself.
With reference to FIG. 10, the piston 21 has a body shaped like a
cup turned upside down slidably mounted within a bushing 22
received in a fluid-tight way within a seat of its own in the body
of the unit associated to each cylinder of the engine.
The slave piston 21 is prearranged for driving the stem 8 of the
respective valve 7 by interposition of a hydraulic tappet 400 (as
already illustrated schematically in FIG. 2).
The tappet 400 has an outer tappet element 400A set within a
widened mouth of the bushing 22, on the outside of the cylindrical
cavity 220 within which the slave piston 21 is slidably guided. The
outer tappet element 400A is slidably mounted on the bottom end of
an inner tappet element 400B. The inner tappet element 400B has a
cylindrical body slidably mounted in the cavity 220 and a top end
in contact with the bottom end (as viewed the drawing) of the
piston 21. The inner tappet element 400B has an internal cavity
that receives pressurized oil from the lubrication circuit of the
engine through a channel 402 formed in the body of the unit 20, and
through chambers 407 defined by circumferential grooves formed in
the inner and outer surfaces of the bushing 22 and through radial
holes 405, 406 formed in the wall of the bushing 22 and in the
element 400B. The pressure of the oil within the element 400B is
lower than the pressure that is set up in the high-pressure volume
C when the master piston is in the active phase.
From the internal cavity to the tappet element 400B, the oil can
pass into the internal chamber 401 defined between the tappet
elements 400A and 400B, through a non-return valve having a ball
open/close element 403 recalled into the closing position by a
spring 404.
Adjacent to the top end of the bushing 22, defined around the
bushing 22 is a circumferential chamber 221 which communicates, by
means of a duct not illustrated, with the high-pressure volume C.
The chamber 221 communicates also with radial holes 222 formed
through the wall of the bushing 22.
In the steps in which the top surface of the slave piston 21 is
below the holes 222, as viewed in the drawing, the chamber 212
within the bushing 22 that faces the piston 21 is in communication
with the pressurized volume through the holes 222 and the
circumferential chamber 221. Consequently, during opening of the
engine valve, the oil pushed by the master piston 16 can enter the
chamber of the slave piston 21 and cause movement thereof, with
consequent movement of opening of the engine valve, via the
hydraulic tappet 400. During closing of the engine valve, the oil
can return into the volume C passing through the same passages.
However, in the final step of the movement of closing of the engine
valve, i.e., when the piston 21 has occluded the holes 222, the
movement of the valve is braked, owing to the fact that the oil
leaving the internal cavity of the bushing 22 is forced to flow
through one or more restricted passages (not visible in FIG. 10)
made in the vicinity of the holes 222 in the bushing 22 (according
to the principle known from the document EP-A-1 344 900 filed in
the name of the present applicant).
In the reverse phase of opening of the valve, during the initial
part of the movement of opening of the valve, the oil coming from
the pressurized volume C can flow only within a chamber 212 above
of the piston 21 passing through a non-return valve 213 carried by
a cap 215 mounted on the top end of the bushing 22. Once the top
surface of the piston 21 has dropped below the level of the holes
222, the oil coming from the pressurized volume C can flow also,
and above all, through the chamber 221 and the holes 222.
The details regarding the slave piston 21 and the hydraulic-braking
device are not in any case described herein any further in so far
as they can be obtained in any one known way and do not fall, taken
in themselves, within the scope of the invention.
According to this embodiment of the invention, integrated within
the known arrangement described with reference to FIG. 10 is a
oscillation dampening device D, illustrated by itself at an
enlarged scale in FIG. 9.
With reference to FIG. 9, the cup-shaped body of the slave piston
21 is used also as body of the oscillation dampening device D. The
cup-shaped body of the piston 21 has an internal cylindrical cavity
211 slidably mounted within which is the movable member D3 of the
oscillation dampening device D, which also has a cup-shaped body.
This movable member D3 is recalled into a resting position against
the bottom wall of the cup-shaped body of the piston 21 by a
helical spring D4 that is axially interposed between the bottom
wall of the movable member D3 and the bottom wall of a further
cup-shaped element D5 rigidly connected to the body of the piston
21. The two cup-shaped bodies of the elements D3 and D5 have their
cavities facing one another in order to receive the spring D4
between them. The helical spring D4 rests against the bottom of the
element D5 preferably via interposition of a spacer ring D12 (the
thickness of which may be chosen as a function of the loading to be
assigned to the spring, which determines the pressure that brings
about triggering of the damper device) and tends to maintain the
movable member D3 in the resting position.
The chamber 212 defined within the piston 21 by the movable element
D3 communicates with the high-pressure volume C via a restricted
opening D10 formed in the bottom wall of the cup-shaped body of the
piston 21. In the case of pressure peaks in the high-pressure
volume C, which lead the pressure to exceed the aforesaid threshold
value, the movable member D3 displaces against the action of the
spring D4, thus creating an additional volume in the space left
free within the cavity 211 by the movable member D3. This
additional volume is, as has been said, in communication with the
high-pressure volume C and consequently causes a simultaneous
increase of the latter in such a way as to dampen the pressure
oscillations, without on the other hand modifying in any
appreciably way the travel imparted on the engine valve. This is
obtained in so far as the characteristics of the spring, its
loading, and the dimensions of the additional volume are
predetermined in such a way as to produce only a dampening of the
pressure peaks of the volume C, when the pressure therein exceeds
the predetermined value.
FIGS. 11 and 11A illustrate a further variant of the system
according to the invention, in which the oscillation dampening
device D is made and integrated in the body of the master piston
16. The master piston 16 has, in a way in itself known, an end
portion 161 designed to receive, directly or indirectly, the thrust
of an actuation cam, and an opposite end portion 162 facing the
high-pressure volume C. In this case, the body of the master piston
16 has a tubular conformation, with an internal cavity 163 rigidly
connected inside which is the body D1 of the oscillation dampening
device D, which in this case is in the form of a cup-shaped element
with an open mouth facing the end 162 that faces the high-pressure
volume C. Slidably mounted within the body D1 is a movable member
D3, which is also cup-shaped and has a bottom wall facing the
high-pressure environment C. On the opposite side of its bottom
wall the movable member D3 is subject to the thrust of a spring D4
that is interposed between the member D3 and a bottom wall D11 of
the cup-shaped body D1. The bottom wall D11 has a central hole D12
that sets the chamber containing the spring D4 in communication
with the internal cavity 163 of the body of the piston 16. The
chamber 163 in turn communicates with the low-pressure environment
of the circuit for supply of the oil through a hole 164 formed in
the wall of the body of the master piston 16 and through the
recirculation line 800 (FIG. 11). As an alternative, in the case
where a dynamic seal constituted by rings made of plastic material
set between the movable member D3 and the body D1 is used, the
communication between the chamber and the low-pressure environment
may be eliminated.
During normal operation of the system, the master piston 16 moves
under the action imparted by the cam, without the movable member D3
moving away from its resting position. However, in the case where
in the high-pressure environment C there arise pressure peaks above
a predetermined threshold value, the plunger D3 moves away from its
resting position, overcoming the action of the spring D4 and
leaving an additional volume inside the piston 16 free, which
causes an attenuation of the pressure oscillations.
As schematically illustrated in FIG. 12, the oscillation dampening
device could also be associated to, and/or integrated in, the
electrically operated control valve 24.
All the embodiments described above envisage use of a device for
dampening pressure oscillations that is designed to intervene
automatically, whenever in the high-pressure volume C there arise
pressure peaks above a predetermined threshold value.
FIG. 13 illustrates a variant in which the device D is of a
controlled type. In this case the device includes an electrically
driven actuator DX (for example, a solenoid, or a piezoelectric
actuator, or a magnetostrictive actuator) designed to cause a
displacement of the movable member D3 that gives rise to a
simultaneous increase of the high-pressure volume in order to
dampen pressure oscillations that are set up in this volume.
The scheme of FIG. 13 may be applied to any embodiment of the
damper device D, for example to any of the embodiments of FIGS. 5A,
5B, 9, 10, 11A, and 12, by providing the aforesaid actuator DX in
order to govern a controlled and desired movement of the movable
member D3.
The actuator DX is controlled by the electronic control unit 25 for
example in a closed-loop mode, on the basis of the signal from one
or more sensors P designed to detect the pressure in the
high-pressure volume C, or else in an open-loop mode, on the basis
of maps stored as a function of the different operating conditions
of the system and/or of the engine.
As already mentioned above, the advantage of a controlled device of
the type illustrated in FIG. 13 lies in the fact that the threshold
value triggering the actuator DX is not always the same as in
self-triggering devices, but rather can be varied according to the
operating conditions. The actuator DX may be of an ON/OFF type or
else of a proportional type.
Naturally, without prejudice to the principle of the invention, the
embodiments and the details of construction may vary widely with
respect to what has described and illustrated herein purely by way
of example, without thereby departing from the scope of the present
invention.
* * * * *